Friction false twist unit

ABSTRACT

The invention relates to a friction false twist unit for crimping synthetic filament yarns, the friction element shafts of which are movable with a radial play, and which are influenced in their movement by friction dampers.

BACKGROUND OF THE INVENTION

The present invention relates to an apparatus for friction falsetwisting an advancing yarn, of the type having a yarn twisting assemblycomposed of at least three spindles mounted for rotation about parallelaxes. More particularly, the present invention relates to a novelmounting structure for the spindles of such an assembly.

Such a friction false twist unit is known, for example, from DE OS 29 36791. In this known friction false twist unit, the driven shafts, alsodescribed as friction element spindles, are supported at their one endwith a radial play in a stationary mounting support or bedplate of thefriction false twist unit, whereas at their other end they are arrangedradially fixed in the bedplate.

Within the radial play, the friction element shaft is flexibly supportedand suspended in a certain manner for damped movement.

In the known embodiment this is accomplished in that the nonrotatingbearing parts which are the outer rings or races of the antifrictionbearings, are supported relative to the stationary mounting supportthrough an interposed rubber ring, which permits on the one hand acertain radial mobility of the driven shaft, and on the other has afunction as a damping element, in that the rubber ring is squeezed inaccordance with the deflection of the flexible bearing.

The friction unit is intended for an operation low in vibrations, andresonant vibrations are to be avoided.

This requirement is also met in the case of the friction unit which isknown from DE OS 29 36 845. In this known friction unit, the frictionelement shafts are likewise supported with a radial play relative to thestationary mounting support in rubber rings, which exhibit dampingproperties.

These rubber rings are practice-proven, since they are reliable andwear-resistant, and moreover ensure a simple assembly of the frictionfalse twist units.

When assembling the known friction false twist units, it is necessary toinsert the rubber rings, together with the bearings, into the boresreceiving the bearings in axial longitudinal direction of the shaftswhile radially compressing them.

However, the possibilities of installing the rubber rings are limited,since it is not possible to either compress them to any desired extentor insert them into the receiving bores in any desired oversize.

Accordingly, the very good damping properties of the known frictionfalse twist units cannot be further improved without adversely affectingtheir assembly.

However, the development of these friction false twist units attempts toreach increasingly higher rotational speeds, so that vibrationalproblems arise therefrom in particular, when the friction element shaftspass through critical speeds.

In particular the passage through critical speeds requires with respectto the improved development of the friction false twist units animproved damping which is technically simple to realize.

In so doing, it should also be considered that the friction false twistunits are equipped to an increasing extent with longer shafts and largermasses, so that a higher residual unbalance results which canpractically not be balanced.

It is therefore the object of the present invention to further improvethe known friction false twist unit such that its vibrational behavioris damped to a greater extent than before, and that yet its assembly isfurther simplified.

Radial deflections of the shaft are to remain small up to highrotational speeds (at least 20,000 rpm), in spite of the facts that afloating bearing is used, that the shafts are longer and equipped withlarger masses than before, that the shafts and the friction disksarranged thereon cannot be prevented from having a residual unbalance,for example as a result of uneven wear, which can practically not bebalanced, and that the shaft diameters are small. What matters inparticular is that the axial position of the shaft relative to thehousing should not change.

SUMMARY OF THE INVENTION

The above and other objects and advantages of the present invention areachieved by the provision of a false twisting apparatus which comprisesa mounting bedplate, and a yarn twisting assembly comprising at leastthree spindles which are mounted to the bedplate for rotation aboutfixed, parallel axes which are positioned at the corner points of anequilateral polygon having a number of sides corresponding to the numberof spindles. A plurality of circular disks are mounted on each spindlefor rotation therewith and with the disks of the spindles overlapping ata point centrally between the spindles and defining an operative yarnpath of travel extending axially therebetween. The apparatus alsoincludes means for concurrently rotating each of the spindles.

In accordance with the present invention, each of the spindles ismounted in a bore in the bedplate, and at least two bearings are mountedin the bore in an axially spaced apart arrangement between the bore andthe associated spindle. At least one of the bearings includes anon-rotating outer ring which includes a contact surface which isperpendicular to the axis of the associated spindle and an outerperipheral surface which is coaxial with the axis of the associatedspindle. A resilient radial damping ring is disposed circumferentiallybetween the outer surface of the outer ring and the bore in thebedplate, and friction damping means is mounted to the bedplate so as tofrictionally engage the contact surface of the outer ring and therebydampen radial movement of the spindle during operation of the apparatus.

From the invention the advantage results that the damping properties canbe varied over wide ranges, without increasing constructionalexpenditure. The solution is characterized in particular in that,although a so-called "flexible" bearing is present, the usualdisadvantage of the flexible bearing is avoided in that the axis ofrotation of the shaft readjusts itself in any operation and at anychange in the operating condition (rotational speed). Rather, with thepresent invention, during the initial operation, each shaft seeks foritself automatically its axis of rotation which does then no longerchange.

This is accomplished in that the radial movement of the shafts, alsodescribed as friction element spindles, is transmitted to the frictiondampers, and that to establish the damping properties, it is onlynecessary to design the friction damper accordingly.

Consequently, it is necessary to arrange the friction dampers such thatthey engage in those regions of the nonrotating bearing parts, whichperform a relative movement with respect to the stationary mountingsupport. They are the outer rings or races of the antifriction bearingsof the friction element shafts. Supported on same are the frictiondampers, the latter extending between the housing and the outer ring.

It is of importance that one end of the friction element shaft beaxially and radially fixed, and that the other end be axially fixed andradially movable. This end is hereafter described as the free end.

It is furthermore of importance that the friction element shafts besupported for floating movement. Accordingly, the head of each frictionelement shaft is freely movable in radial direction and thereforeperforms deflections. These deflections are excited by periodicalvibrations which need to be considerably damped, in particular whenpassing through critical speeds, so that the friction elements are notdestroyed.

Still further, it is of importance that the friction element shafts befixed absolutely rigidly in axial direction. This is accomplished by astop, against which the bearing of the friction element shaft issupported.

In this arrangement, it matters that not only the installation isaxially fixed, but that also the axial position of the shafts relativeto one another is accurate up to few one hundredths of a millimeter, soas to avoid a contacting of the friction disks, and thus damage to theyarn. The axial biasing may occur in force-locking manner by springelements, for example, metal springs or rings. The axial biasing of bothbearings, however, may also be generated in that the outer ring of theone bearing is secured in its bush under an initial stress, for exampleglued or connected in any other manner. Finally, it is possible to biasthe bearings in that from the beginning the ball tracks of the twobearings are introduced relative to one another into a common outer ringsuch that the bearings are biased relative to one another.

For the invention the mounting support of the shafts is of specialimportance.

This mounting support consists of a fixed bearing and a loose bearing,which meet with the following properties:

As regards the fixed bearing, the outer ring is secured substantiallyaxially relative to the housing. The shaft is likewise securedsubstantially axially relative to the outer ring. The locking isoperative in axial direction. The axial locking may occur, for example,by a mutual biasing of the two antifriction bearings.

It may be useful to make the bearing flexible in the radial direction.In accordance with the invention, this measure allows to influence thevibrational behavior without the shaft being able to move in axialdirection.

The axial locking serves to prevent a contacting of the friction diskswhich form between each other a small spacing for the advance of thesynthetic yarn to be crimped.

The radially flexible bearing is not absolutely necessary, but mayadditionally be needed so as to be able to influence the vibrationalbehavior in a damping manner.

In comparison therewith, in a loose bearing the shaft is axially securedrelative to the outer ring, whereas the outer ring is axially movablerelative to the housing. As a result, the outer ring of the loosebearing is damped in radial direction under the influence of externalcontact pressures of the friction damper, while an accurately definedaxial bearing positioning is obtained for the shaft at the same time.

Since the shaft is likewise secured in axial direction relative to theouter ring, the latter is axially movable relative to the housing, so asto avoid locked-up stress, for example, by thermal expansion, in such amanner that the shaft clearly secured in axial direction is unable toevade uncontrolled movement in the axial direction.

This kind of bearing is realized in accordance with the invention whenthe bearing, while axially secured, enables as a whole in addition ashaft suspension flexibly movable in radial direction.

As has been recognized by the invention, this measure allows to enablean influencing of the vibrational behavior with a flexible and dampedradial movement.

More extensive embodiments avail themselves of a bearing, in which twoindividual antifriction bearings are braced against one another, one ofthe antifriction bearing being the fixed bearing and the other the loosebearing.

As a result of the axial biasing, axially uncontrolled movements of thefriction element shafts are excluded. Therefore, it is avoided withcertainty that the individual friction elements collide with one anotherand are damaged.

The invention makes use of the recognition that the radial vibrations ofthe shafts of a friction unit which arise from unbalance and areactuated periodically, can be damped by friction dampers which aresecured by external contact pressures.

The radial movements of the friction element shafts are thus transmittedby external contact pressures to the friction dampers and influenced bythe friction dampers such that an amplitude increase of vibrations ofthe system is avoided.

Basically, damping elements with a path-dependent restoring force, suchas for example radially biased O-rings, are considered as frictiondampers. In this case, the restoring force is dependent on thedeflection or on the deflection-dependent deformation of the dampingelement.

Alternatively or additionally, also damping elements are used, whichgenerate a path-independent frictional force. In this case, the dampingelements are biased by a normal force in axial direction, whereby thedamping effect is dependent on the amount of the normal force and thecoefficient of friction between the damping element and contact surface.

Special attention is attributed in particular to the last-mentionedfurther development of the invention, since the damping effect may bepredetermined via the amount of the normal force.

Thus, for example, it may be provided that the normal force ispredetermined so high that only during a first acceleration of thefriction unit a one-time breakaway of the damping element with apath-independent frictional force occurs, whereby the friction elementshaft centers itself.

The present invention recognizes that the friction element shaftoperating under unbalance attempts to center itself at the initialstartup. The forces of unbalance occurring in this instance are so greatthat the forces of static friction between the friction dampers and thehousing are overcome. In so doing, the assembled position of theantifriction bearings breaks away, and the shaft shifts to itsself-centered position, until the forces of unbalance are no longercapable of overcoming the forces of sliding friction. Then, the shaftoperates in its self-centered position and remains in same, since it isnow again held by the forces of static friction. In this position, theshaft is figuratively frozen, with the frozen position inducing theslightest vibrational excitations.

It is therefore suggested that the friction dampers be adjusted suchthat the axial, external contact pressures are overcome only when theforces of unbalance engaging on the shafts as a result of unbalancesexceed the value which is necessary to overcome the static frictionbetween the friction damper and the front surface of the outer ring.This will be relevant in the range of dangerous resonant vibrations,when the friction element shafts pass through their critical speeds.

This further development of the invention also has the advantage of avery stiff damping. Consequently, the damping forces are very high andthe deflections very small.

Additionally, the friction dampers of the present invention offer theadvantage that they act statically in the range of the operating speeds.In this speed range, the friction element shaft rotates overcritically.

It is a further characteristic of the invention that the axial lockingof the friction element shafts effects a harder mounting support. As aresult, the vibrational energy of the shafts can be received and dampedonly by a single rubber ring. This rubber ring is arranged facing thefree shaft end and causes by its squeezing a damping of the shaft.However, since contrary to previous solutions only a single rubber ringis present, it is necessary that the friction damper of the presentinvention absorb and destroy the excess of vibrational energy.

The loose bearing preferably is nearest the free end of the shaft, i.e.nearest the disks on the associated shaft, which provides an advantagein that the introduction of torque of the drive into the shafts occursat the fixed end, and that the assembly of the friction unit is adaptedto the drive without influence on the axial locking of the shafts.

The contact pressure against the outer ring of the loose bearing ispreferably operative in the sense of an axial clamping of the loosebearing against the fixed bearing. This offers the advantage that theaxial, external contact pressures do not only introduce the radialmovement of the shaft into the friction damper, but moreover also assumethe axial bracing of the two antifriction bearings. In so doing, theamount of the axial biasing force is to be selected such that thefriction damper fits tightly on the front surface of the outer ring ofthe loose bearing.

It may here be useful that the respective contact pressure required todamp the radial movement of the friction element shafts is adjusted viathe axial biasing force, so that the required labor of assembling can befurther reduced. In this case it is recommended that the frictiondampers be arranged such that on the one hand they support themselves onthe housing, and that on the other hand they support themselves onannular steps or annular offsets, which are directly formed by the outerrings. The direct connection between the outer rings and the frictiondampers allows to accomplish that the variance of the vibratory system,which forms the friction false twist unit, is minimized. This alsoallows to reduce the complex vibrational behavior to a low number ofinfluential parameters.

The friction dampers may take the form of annular cup springs which aresupported between the outer rings of the loose bearings and thebedplate. This provides the advantage that standardized components areused as friction dampers. This further development of the inventionavails itself of the recognition that the radial movement of thefriction element shafts is only small, and that the supporting forces ofthe cup springs can easily be absorbed by their inner or outer edges,and the force is transmitted on the contact points between the inner orouter edge of the cup springs and the outer ring or housing.

In so doing, the necessary force of static friction which transmits theradial vibrational movement of the rotating shafts to the frictiondampers, is generated via the supporting forces.

The initial biasing forces which are applied by means of the cupsprings, can be adapted over wide ranges to the respective case ofapplication, since the characteristic curves of the cup springs can beinfluenced by parallel or serial arrangement. Furthermore, a pluralityof cup springs with identical dimensions, but different springstiffnesses is available, so that it is possible to appropriately selecteach time the optimal cup spring for each application.

It is known, though, from DE OS 29 36 845 to mutually brace the bearingsof a friction element shaft by means of cup springs. However, this caseof application differs substantially from the invention in that the cupspring in the known construction serves to prestress the bearing onlyaxially. Furthermore, the cup spring is installed at the end of thefriction element shaft which is firmly secured in radial direction.

In the case of the invention, however, it matters that the free end ofthe friction element shaft may move radially, since only this allows theaction of the forces of static friction to come to bear. Thesefrictional forces cause a deformation of the cup spring in radialdirection. In particular, cup springs permit a deformation with a greatdamping portion to occur in radial direction, which is used by thepresent invention.

The outer rings of the fixed bearings may be secured relative to thebedplate by axially biased rubber rings, which are clamped between theouter rings and the bedplate. Thus the fixed shaft end is secured inposition by means of a rubber ring. The amount of the clamping forceresults from the respective contact pressure, at which the rubber ringis axially compressed. To this end, a cup spring is used in particularin addition to the rubber ring, which biases the rubber ring against astepped diameter or collar of the outer ring of the antifrictionbearing. The cooperation of rubber ring and cup spring allows on the onehand to make use of the punctual flexibility of the rubber ring, whereasthe amount of the axial clamping force is defined by the cup spring. Inso proceeding, it should be noted the axial clamping force is so greatthat the clamped outer ring of the bearing is unable to lift against theforce of the cup spring.

The further development of the invention offers the advantage ofsimplest manufacture and assembly. In this embodiment the outer ring isaxially secured, for example, by a clamping ring which holds the outerring against the pressure plate.

In a further embodiment of the invention, the outer rings of the fixedbearings may be clamped between a pressure plate and a counter pressureplate of the bedplate. This embodiment offers a simple possibility ofaxially bracing two individual antifriction bearings. To this end, theouter ring at the free end of the shaft is initially accommodated forsliding movement in the outer ring of the antifriction bearing at thefixed end of the shaft. It is then held, for example by a cup spring,relative to the outer ring of the fixed shaft end under an axial bias,and secured in this position, as is described with reference to anembodiment.

The friction dampers of the present invention are preferably operativesubstantially in the region of the free ends of the shafts. Thisprovides the advantage that a favorable force introduction point for thestatic friction forces is made available, so that high damping effectsare obtained by means of relatively little contact pressures. In thisinstance, the effective damping force of the friction damper acts uponthe shaft as a moment, with the length of the lever arm between theforce introduction point into the shaft and the fixed bearing assumingits greatest possible value.

The shafts may be flexibly supported in the bore of the bedplate bymeans of annular radial rubber dampers, such that the rubber dampers arecompressed upon a radial movement of the shaft. This feature recognizesthat the annular rubber dampers can assist in centering the frictionelement shafts.

Such annular rubber dampers, as is known, are able to assist a fastrotating, vertically oriented shaft in its attempt to self-center.

It should explicitly be noted that this advantage is not limited only todampers which consist of rubber, but basically the dampers may alsoconsist of any comparable elastomer.

The antifriction bearings of each shaft may be biased in the axialdirection toward each other, so as to eliminate play. This permits theshaft deflections to be transmitted entirely free of play to the outerrings of the bearing, and to thus damp same already at their formation.

BRIEF DESCRIPTION OF THE DRAWINGS

Some of the objects and advantages of the present invention having beenstated, others will appear as the description proceeds, when taken inconjunction with the accompanying drawings, in which:

FIG. 1 illustrates a first embodiment of the present invention whichemploys a single friction element shaft with a fixed shaft end;

FIG. 1a illustrates a further development of the embodiment of FIG. 1with a clamped shaft end;

FIG. 2 illustrates a second embodiment of the present invention whichemploys a single friction element shaft;

FIG. 2a shows a detail view of the embodiment of FIG. 2 to illustratethe bearing at the free shaft end;

FIG. 2b shows a detail view of an alternative bearing of the embodimentof FIG. 2 at the fixed shaft end;

FIG. 3 is an axial top view of a friction false twist unit with threefriction element shafts;

FIG. 4 illustrates an embodiment of the friction dampers in accordancewith the invention installed on a friction element shaft;

FIG. 5 is a diagram illustrating the radial deflections as a function ofthe rotational speed; and

FIGS. 6a and 6b show each an embodiment of the present invention inaccordance with FIGS. 1-4 with perforce centered cup springs.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Unless specified otherwise in the following, the description will alwaysapply to all FIGS. 1-4 and 6a-b.

FIG. 3 is an axial top view of a friction false twist unit 1 forcrimping synthetic fibers. In this illustration, the friction falsetwist unit comprises three vertically standing friction element shafts2, which are arranged in the axial top view of FIG. 3 in the form of anequilateral triangle such that friction elements 8, which are in thepresent embodiment constructed as friction disks, overlap in the centerof the triangle. In this respect, reference may be made to the entirecontents of DE-OS 29 36 791 and DE-OS 29 36 845.

As is further shown in FIGS. 1, 1a, 2, 2a-b, and 4, the friction elementshafts 2 are supported relative to a stationary mounting support 3, alsodescribed as housing 3, for rotation in antifriction bearings 4. In thecase of the embodiments of FIGS. 1, 1a, 2, 2b, and 4, the stationaryhousing 3 is closed at its front ends by a pressure plate 14 and acounterpressure plate 14.1.

Each of bearings 4 consists of a peripheral groove or track notdescribed in more detail, which rotates along with friction elementshaft 2 and a nonrotating bearing part 4.1 generally termed as outerring or race 4.1.

Outer ring 4.1 is rotationally immobile with respect to the housing. Theouter rings 4.1 at free shaft end 7 exhibit with respect to housing 3 aradial play 5, which is dimensioned such that friction element shaft 2is able to perform a movement caused by unbalance, without contactinghousing 3. With respect thereto, reference is made to the entirecontents of the aforesaid prior art.

This unbalance-generated movement of outer rings 4.1 can be damped in amanner known per se, in that the outer rings are arranged within theradial play in radial rubber dampers 17.1 and 17.2 which again aresupported on housing 3.

However, it should be stated explicitly that this does not represent alimitation of the invention, but that the invention can also be realizedwithout these radial rubber dampers 17.1 and 17.2 respectively.

Common to all embodiments is that one end of the rotating shafts isfixed with respect to housing 3, for example, end 6 in FIG. 4, and thatthe opposite end is freely movable within radial play 5.

A special characteristic exists in FIG. 1.

Although, here, both shaft ends are shown movable (=free shaft end 7),same applies only to the instance of the initial startup of the frictionfalse twist unit. At this occasion, the rotating shaft centers itselfwithin housing 3 in its self-centered position, a radial sliding of thelower shaft end being basically not excluded either.

However, as soon as the shaft has assumed this self-centered position,it freezes in this position and the then slight forces of unbalance willno longer suffice to break away the lower shaft end from such positionagain.

In this condition the lower shaft end acts statically, i.e., in thefashion of a fixed shaft end.

The secured shaft end 6 is positioned relative to housing 3 radiallyimmovably in lower bearing 4.

As further shown in FIGS. 1, 1a, 2, 2a-b, and 4, arranged between outerrings 4.1 and housing 3 or pressure plates 14 or 14.1 are frictiondampers 9, which counteract the radial movement of friction elementshafts 2 within radial play 5 by external friction forces.

These external friction forces are transmitted via paired frictionelements 15 from the unbalance-generated vibrational movement offriction element shaft 2 to stationary mounting support 3. This actionis done by static friction, the further function of which is describedbelow in greater detail.

This effect is accomplished in that outer ring 4.1 of the loose bearingis acted upon by a friction damper 9, which is adjusted by axialexternal contact pressures against an end surface of the outer ring, andwhich damps the radial movement of shaft 2.

Except the instance of the lower portion of FIG. 2, the friction dampers9 consist exclusively of annular cup springs 10 which are peripherallysupported between outer rings 4.1 and stationary mounting support(housing) 3 or pressure plates 14 or 14.1.

At this point however, it should explicitly be pointed out that this isnot to limit the invention to cup springs.

It is possible to use all types of friction dampers which are arrangedbetween outer rings 4.1 and stationary mounting support (housing) 3, andtheir use allows to transmit the unbalance-generated movement of thefriction element shaft by external static friction forces to a frictiondamper and to thus damp same.

The use of cup springs 10 allows to accomplish in a very simple mannerthat outer rings 4.1 are secured in position relative to stationarymounting support 3, 14, 14.1 by friction dampers 9 under an axialbiasing force such that the force of friction is dependent on therespective biasing force.

This offers the advantage that the amount of the damping force offriction can be predetermined by the selection of a cup spring 10 with acorresponding characteristic curve.

Thus, it is possible in the instance of small dimensions of installationto produce nonetheless high damping forces by selecting acorrespondingly stiff cup spring, should need arise.

A possibility of firmly securing a shaft end in position is shown in thelower portion of FIG. 2.

There, a counterpressure cup spring 10.1 is supported between stationarycounterpressure plate 14.1, an interposed axial ring 11, and a pressurering 12 on a collar 13, which forms part of outer ring 4.1.

The annular pressure ring 12 which is an axially biased rubber element,is arranged and held in position on a stepped diameter 13.1 betweenaxial ring 11 and collar

In this instance, an axial biasing is applied by counterpressure cupspring 10.1 which pushes pressure ring 12 under axial stress fromcounterpressure plate 14.1, via centering sleeve 16, toward pressureplate 14.

In this instance, counterpressure cup spring 10.1 is stronger than cupspring 10, so as to produce high frictional forces necessary to retainthe fixed bearing located at this shaft end, between pressure ring 12and axial ring 11 or collar 13.

Furthermore, in this manner it is avoided with certainty that thecentering sleeve rises from pressure plate 14.

In a general sense, this also applies to the embodiment shown in FIG.2b, however with the exception that a contact exists betweencounterpressure plate 14.1 and housing 3. Thus, a precisely definedinstallation position is obtained between counterpressure plate 14.1 andhousing 3, the initial bias of counterpressure cup spring 10.1, as inthe case of FIG. 2, being large enough so as to axially and radiallysecure the respective shaft end in the meaning of a fixed bearing.

Furthermore, in this embodiment pressure ring 12 assumes a centeringfunction for the friction element shaft already during the assembly. Tothis end, counterpressure ring 12 is supported with its outercircumference on housing 3 and with its inner circumference on anextension of outer ring 4.1 of the bearing.

A further characteristic consists in that in this instancecounterpressure cup spring 10.1 is perforce centered with its outsidediameter in housing 3. To this end, the outside diameter ofcounterpressure cup spring 10.1 corresponds to the inside diameter ofhousing 3.

Applicable to all Figures is that the upper antifriction bearing and thelower antifriction bearing 4 are biased against one another in axialdirection.

This biasing of the bearings serves the purpose of transmitting theshaft movement free of play to the outer rings of the bearings and thedamping elements being active thereon. This free-of-play transmission isaccomplished in that the damping action becomes fully effective alreadyat the slightest deflections.

In the case of FIGS. 2 and 2b, the axial biasing of the bearings occursvia common outer ring 4.1, for example, by predetermined, slightlydiffering spacings of the ball tracks in the outer ring with respect tothe ball tracks in the inner ring.

Another possibility of applying a bias between the bearings may occur inthe instance of a common outer ring of the two antifriction bearings byinserting balls with a slightly oversized diameter.

In all embodiments, friction dampers 9 are operative substantially inthe regions of the free ends 7 of friction elements shafts 2.

In the cases of FIGS. 1, 1a, 2a-b, and 4, the friction element shafts 2are arranged in addition in annular, radial rubber dampers 17.1, 17.2(or 12). The radial rubber dampers 17.1 serve as upper centering ring offriction element shafts 2, whereas radial rubber damper 17.2 (or 12)serves analogously as lower centering ring, so as to damp the shaft atthe initial startup, before the latter assumes its self-centeredposition.

In all cases, a radial compression of radial rubber dampers 17.1 and17.2 (or 12) occurs, when friction element shaft 2 performs, as itrotates, an unbalance-generated deflection from its illustrated centralposition.

As can be noted from FIG. 2, the friction element shafts 2 are set intorotation by a drive motor 18, which drives all friction element shafts 2(see FIG. 3) via a toothed belt not shown, which is operative on toothedbelt pulleys 19.

FIG. 1a shows a characteristic of the invention. In this case, the outerring of the lower bearing is secured from the beginning in radialdirection. This is accomplished in that the outer ring is provided atits lower end with a collet, which clamps the outer ring betweencounterpressure plate 14.1 and the lower front end of centering sleeve16. The centering sleeve 16 is supported with its upper end on pressureplate 14. Thus, it is located immovably between pressure plate 14 andcounterpressure plate 14.1 as is the outer ring of the lower bearing. Ofimportance to this end is that counterpressure plate 14.1 can be clampedfirmly against housing 3 by means of clamping screws 23. To this end, anair gap having the dimensions of the collet is provided betweencounterpressure plate 14.1 and the lower front end of housing 3.

The outer ring of upper antifriction bearing 4 is pressed downward bycup spring 10, the bias applied by the latter being so great that theentire shaft is secured in axial direction. In this instance, the cupspring thus assumes the double function of axially securing the shaft onthe one hand and of damping the vibrations of the shaft on the other.Both functions are dependent on one another via the biasing force andthe coefficient of friction.

FIG. 2a shows a further embodiment of axially securing the antifrictionbearings.

The illustration of FIG. 2a can be a cutaway portion of FIG. 2. Asregards all particulars not described in more detail, reference may bemade to this Figure and its relevant description in their entirety.

Essential is that outer ring 4.1 of the upper ball bearing isaccommodated in a bearing seat which is formed in the outer ring of thelower ball bearing. Initially, the outer ring of the upper ball bearingis arranged for sliding movement in the bearing seat, an axial biasingspring 26 being located between an annular offset of the bearing seatand the front end of the outer ring of the upper antifriction bearingfacing the insert side. The outer ring 4.1 of the upper antifrictionbearing is pressed into the bearing seat against the force of axialbiasing spring 26, until the necessary biasing force is reached. In thisposition, outer ring 4.1 of the upper antifriction bearing is secured.To this end, it is provided on its outer circumference with an annulargroove 25 which communicates with a bore 24. The latter extends throughthe outer ring of the lower bearing and terminates outside. To securethe outer ring of the upper bearing, after applying the axial bias,annular groove 25 is filled through bore 24 with an adhesive whichsecures the two bearings in an axially clamped position.

Operation:

In all embodiments, friction element shafts 2 are arranged relative tohousing 3, 14, or 14.1 with radial play 5. As can be noted, the radialplay 5 occurs only at the free shaft end.

As is known, such friction element shafts 2 rotate at speeds up to25,000 revolutions per minute and greater. In so doing, rotating forcesare caused by always present unbalances, which are contributory to ageneration of vibrations of the entire arrangement.

The vibration of the arrangement is thus caused by periodicallyoccurring forces of unbalance which can be counteracted by oppositelydirected damping forces.

These damping forces are generated by friction dampers 9, since they areconnected on the one hand with stationary mounting support 3 or 14 or14.1 and on the other hand with periodically moved outer rings 4.1 ofthe loose bearings. Thus, at the points of contact between the frictiondampers and the surfaces moved relative thereto a friction force occurs,the amount of which is dependent on the amount of the force of contacton the friction surface and paired friction elements 15. Suitably, theamount of the contact force is to be selected such that upon exceeding apredetermined maximum force, the paired friction element breaks away anda condition of sliding friction is reached, so that the rotating shaftis able to center itself.

Thereafter, the friction dampers operate statically, i.e., a renewedbreakaway will no longer occur under normal conditions.

Furthermore, the friction dampers of the present invention allow toobtain a damping effect also during the acceleration of the frictionunit, when the latter is brought from standstill to its operating speedand, in so doing, has to pass through the particularly dangerous naturalresonant frequencies of the lower critical speeds.

After the friction element shaft has passed through its critical speed,occurring vibrational deflections fade away and thus also the movementsof the outer ring. In the final condition, with appropriately selectedpaired friction surfaces, the friction damper will assume a staticcondition, since in this instance the shaft rotates self-centered.

Shown in FIG. 5 is a diagram, in which the dynamic radial deflections ofa friction element shaft in accordance with the invention are plottedqualitatively as a function of the disk speed. In this diagram, thecourse of curve I corresponds to a qualitative course, as can beobtained with a standard flexible bearing of the friction element shaft.This course is illustrated, so as to be able to show the effect of thepresent invention with reference to the course of curve II. The latterhas as a whole a flatter slope, with lesser deflections as a whole(approximately up to two thirds less), as can be obtained with a bearingof the present invention.

As the speed increases, in both cases the deflections start to increaseand reach at about 8,000 rpm their maximum, which is however in case IIby about two thirds less than in case I.

Accordingly, in the range of this speed, the first critical speed of thefriction element shaft is reached, at which unbalance-generated radialdeflections would assume dangerously high values, were they notcounteracted by the friction dampers of the present invention.

Thereafter, in case of curve I two further maxima occur in the range of13,000 rpm and 22,000 rpm, to which the foregoing applies accordingly,whereas in case of curve II, after passing the first critical speed, theradial deflections decrease continuously until the desired speed of thefriction element shaft of about 24,000 rpm is reached.

Furthermore, it is striking that in case of curve I three criticalranges of speeds are passed, whereas with the bearing of case of curveII only a single range of critical speed occurs.

As is further shown in FIGS. 1, 1a, 2, 2a-b, and 4, the cup springs 10are arranged in a centered position with respect to the axis center. Inthis instance, the edges of cup springs 10 form a spacing at a constantannular width over the circumference both with respect to the frictionelement shaft and with respect to centering sleeve 16.

For certain cases of applications it may moreover be advantageous toforcibly center the cup springs 10 at least with one end.

To this end, FIGS. 6a-b show that the cup springs are arranged relativeto housing 3 inside a centering ring 27, the inside diameter of whichcorresponds exactly with the outside diameter of the cup spring.

I claim:
 1. An apparatus for friction false twisting an advancing yarnand comprisinga mounting bedplate, said bedplate including a pressureplate and a counter pressure plate which are axially spaced apart andparallel to each other, a yarn twisting assembly comprising at leastthree spindles, and means mounting said spindles to said bedplate forrotation about fixed, parallel axes which are positioned at the cornerpoints of an equilateral polygon having a number of sides correspondingto the number of spindles, and a plurality of circular disks mounted oneach spindle for rotation therewith, and with the disks of the spindlesoverlapping at a point centrally between said spindles and defining anoperative yarn path of travel extending axially therebetween, and meansfor concurrently rotating each of said spindles, said mounting meansincluding for each of said spindles a bore in said bedplate, at leasttwo bearings mounted in said bore in an axially spaced apart arrangementbetween said bore and the associated spindle, with the one of saidbearings located nearest the circular disks of the associated spindleincluding a non-rotating outer ring which includes a contact surfacewhich is perpendicular to the axis of the associated spindle and facesthe circular disks of the associated spindle, and an outer peripheralsurface which is coaxial with the axis of the associated spindle, andthe other of said at least two bearings including a non-rotating outerring which is clamped between said pressure plate and said counterpressure plate of said bedplate, a resilient radial damping ringdisposed circumferentially between said outer peripheral surface of saidouter ring of said one bearing and the outer ring of said other bearing,and friction damping means mounted to said bedplate so as tofrictionally engage said contact surface of said outer ring of said onebearing and bias the same toward said other bearing and to therebydampen radial movement of the spindle during operation of saidapparatus.
 2. The apparatus as defined in claim 1 wherein said frictiondamping means comprises an annular cup spring which is positionedbetween said contact surface and one of said pressure plates of saidbedplate.
 3. The apparatus as defined in claim 1 wherein said outer ringof said other bearing includes an annular bearing seat which receivesthe outer ring of said one bearing therein.
 4. The apparatus as definedin claim 3 wherein said friction damping means comprises an annular cupspring which is positioned between said contact surface and one of saidpressure plates.
 5. The apparatus as defined in claim 4 wherein saidresilient radial damping ring is disposed between said outer peripheralsurface of said outer ring of said one bearing and said bearing seat ofsaid outer ring of said other bearing.